Cooling device for electronic components

ABSTRACT

A cooling device ( 10 ) including a chamber having an evaporator section ( 34 ) and a condenser section, the condenser section extending from a periphery of the evaporator section ( 34 ).

FIELD OF THE INVENTION

The present invention relates to cooling devices for electroniccomponents.

BACKGROUND OF THE INVENTION

Advances in photolithography have enabled dense packing of transistorson a chip. This has in turn resulted in an increase in the amount ofheat generated per unit area of the chip, in some instances, exceedingthe thermal design limit of conventional cooling devices. There is thusa need for a cooling device that is capable of dissipating the wasteheat generated by electronic components effectively.

SUMMARY OF THE INVENTION

In a first aspect, the invention provides a cooling device comprising achamber having an evaporator section and a condenser section, thecondenser section extending from a periphery of the evaporator section.

A compact cooling device is thus achieved with the back-to-backconfiguration of the evaporator and condenser sections. The compactarrangement of the cooling device results in almost zero pressuredifferential between the evaporator and condenser sections (and hence, auniform temperature distribution in the chamber) for fluid flow withinthe chamber. This reduces the overall thermal resistance of the coolingdevice.

A working fluid may be arranged to absorb and transfer heat energy. Theworking fluid may have a latent heat of vaporisation of greater than orequal to about 1500 joules per gram (J/g). The working fluid may includea plurality of nanoparticles, the nanoparticles having a thermalconductivity of between about 200 Watt per metre Kelvin (W/m·K) to about400 W/m·K.

The condenser section may comprise a plurality of condenser finsextending from the periphery of the evaporator section.

A plurality of convective fins may extend from surfaces of the condenserfins. This facilitates heat transfer from the condenser fins.

The convective fins may extend between adjacent ones of the condenserfins. This enhances the compactness of the cooling device.

A fan may be arranged to direct a flow of air at the convective fins.This enhances convective heat transfer from the convective fins and thusimproves the heat rejection rate.

An air guide may be arranged to direct the flow of air from the fan tothe convective fins. This enhances convective heat transfer from theconvective fins.

The convective fins may be integrally moulded to the surfaces of thecondenser fins. This reduces thermal contact resistance and thusincreases thermal conductivity between the convective fins and thecondenser fins.

A capillary element may be arranged to facilitate return of condensatefrom the condenser section to the evaporator section. This facilitatesreturn of the condensate to the evaporator section regardless of theorientation of the cooling device.

A base portion of the condenser section may be at an angle to a baseportion of the evaporator section. This facilitates return of thecondensate to the evaporator section by gravitational means.

The evaporator section may have a capillary structure. The capillarystructure enhances boiling in the evaporator section as it creates amicro-flow situation within the evaporator section by drawing heatedworking fluid away from the base portion of the evaporator section andfresh working fluid in the form of condensate from the condenser fins tothe evaporator section. The degree of superheat (i.e. the differencebetween the temperature of the base portion of the evaporator sectionand the saturation temperature of the working fluid) is thus reduced.The capillary structure may comprise an open cell foam metal. The opencell foam metal may be bonded to a base portion of the evaporatorsection by diffused bonding or brazing to reduce thermal contactresistance and thus improve thermal conductivity between the baseportion of the evaporator section and the capillary structure. Thecapillary structure may have a pore density of from about 8 to about 52pores per centimetre (ppc) and more particularly, a pore density ofabout 16 ppc.

Other aspects and advantages of the invention will become apparent fromthe following detailed description, taken in conjunction with theaccompanying drawings, illustrating by way of example the principles ofthe invention.

BRIEF DESCRIPTION OF THE DRAWINGS

Embodiments of the invention will now be described, by way of exampleonly, with reference to the accompanying drawings, in which:

FIG. 1A is an exploded top perspective view of a cooling device inaccordance with one embodiment of the present invention;

FIG. 1B is an exploded bottom perspective view of the cooling device ofFIG. 1A;

FIG. 2A is an assembled top perspective view of the cooling device ofFIG. 1A;

FIG. 2B is an assembled bottom perspective view of the cooling device ofFIG. 1B;

FIG. 3 is a schematic cross-sectional view of the cooling device takenalong section lines A-B and B-C in FIG. 2B; and

FIG. 4 is a graph showing the transient performance of a cooling devicewith convective fin thickness of 0.25 millimetres (mm).

DETAILED DESCRIPTION OF EXEMPLARY EMBODIMENTS

The detailed description set forth below in connection with the appendeddrawings is intended as a description of the presently preferredembodiments of the invention, and is not intended to represent the onlyform in which the present invention may be practiced. It is to beunderstood that the same or equivalent functions may be accomplished bydifferent embodiments that are intended to be encompassed within thescope of the invention.

A cooling device 10 for an electronic component (not shown) such as, forexample, a microprocessor or central processing unit (CPU) of a computerwill now be described below with reference to FIGS. 1A, 1B, 2A, 2B and3.

The cooling device 10 comprises a fan 12 mounted to an air guide 14, ahermetically sealed chamber 16 having a plurality of convective fins 18and a sleeve 20, and an attachment member 22 in the form of a legsupport. The hermetically sealed chamber 16 is provided with a workingfluid 24 to absorb and transfer heat energy.

The fan 12 is arranged to direct a flow of air at the convective fins 18and the air guide 14 is arranged to direct the flow of air from the fan12 to the convective fins 18. In the present embodiment, the fan 12 ismounted to the air guide 14 by a plurality of beams 26 extending fromthe stator 28 of the fan 12 to the air guide 14, the fan 12 beingpositioned such that a separation is maintained between the rotor 30 ofthe fan 12 and the hermetically sealed chamber 16. As can be seen fromFIGS. 2B and 3, the air guide 14 is spaced from the outermost convectivefin 18 to allow airflow through.

As shown in FIG. 3, the hermetically sealed chamber 16 is mounted to theair guide 14 via a clipping mechanism 32. The hermetically sealedchamber 16 comprises an evaporator section 34 and a condenser sectioncomprising a plurality of condenser fins 36 extending from a peripheryof the evaporator section 34. In the present embodiment, a base portionof the condenser section is slanted at an angle θ to a base portion ofthe evaporator section 34. This facilitates return of condensate 38 fromthe condenser fins 36 to the evaporator section 34 by gravitationalmeans. The base portion of the evaporator section 34 in contact with theelectronic component may be polished to reduce thermal contactresistance and thus improve heat transmission from the electroniccomponent to the evaporator section 34.

In the present embodiment, the evaporator section 34 is provided with acapillary structure 40. The capillary structure 40 may, for example, beprovided in the form of an open cell foam metal made of copper;aluminium or other highly thermally conductive material and may bebonded to the base portion of the evaporator section 34 by, for example,diffused bonding or brazing to reduce thermal contact resistance andthus improve thermal conductivity between the base portion of theevaporator section 34 and the capillary structure 40. The capillarystructure 40 may have a pore density of from about 8 to about 52 poresper centimetre (ppc) (equivalent to from about 20 to about 130 pores perinch (ppi)) and more particularly, a pore density of about 16 ppc.

The condenser fins 36 are hollow and of a certain thickness with one endopen towards the evaporator section 34 to allow working fluid 24 in thevapour phase to enter from the evaporator section 34 and condense evenon its outermost end, thereby spreading heat effectively and achieving amore uniform temperature distribution. In the present embodiment, eachof the condenser fins 36 is provided with a capillary element 42 suchas, for example, a cotton thread or braided copper wires to facilitatereturn of the condensate 38 from the condenser section to the evaporatorsection 34 at various orientations of the cooling device 10. A smallgroove (not shown) may be provided on the slanted bottom of thecondenser fin 36 for receiving the capillary element 42 in the form of acapillary pumping thread.

The working fluid 24 in the hermetically sealed chamber 16 may be asingle or multi component condensable fluid. Examples of the workingfluid 24 include, but are not limited to, water, dielectric fluids suchas, for example, hydrofluoroethane (HFE) or fluorocarbon (FC) in therefrigerant series, or any other fluid or combination of fluids with ahigh latent heat of evaporation (more particularly, a latent heat ofvaporisation of greater than or equal to about 1500 joules per gram(J/g)) and which can permeate the foam metal material in the evaporatorsection 34. The working fluid 24 may be provided with a plurality ofnanoparticles to augment the effective thermal conductivity of theworking fluid 24. The nanoparticles may have a thermal conductivity ofbetween about 200 Watt per metre Kelvin (W/m·K) to about 400 W/m·K. Anoptimum amount of working fluid 24 is provided in the hermeticallysealed chamber 16 such that the capillary structure 40 is always incontact with and saturated with the working fluid 24 to prevent dryingout and to ensure wetting of the capillary structure 40 throughout theoperation of the cooling device 10. The optimum amount of working fluid24 in the hermetically sealed chamber 16 is determined based on theamount of heat generated by the electronic component. A larger quantityof working fluid 24 is required in instances where more heat isgenerated to prevent dry out. The working fluid 24 in the hermeticallysealed chamber 16 may be in its liquid-vapour saturated state tofacilitate conversion of the working fluid 24 into its vapour form asthis enhances heat dissipation by the cooling device 10. To achieve theliquid-vapour saturated state, the hermetically sealed chamber 16 may bein a pressurized or vacuum state depending on the working fluid 24employed. In other words, the pressure in the hermetically sealedchamber 16 is dependent on the saturation properties of the workingfluid 24. As an example, the hermetically sealed chamber 16 may be in avacuum state when water is employed as the working fluid 24.

As can be seen from FIGS. 1A, 1B, 2A and 2B, the convective fins 18extend from surfaces of the condenser fins 36 and between adjacent onesof the condenser fins 36. The convective fins 18 may be integrallymoulded to the surfaces of the condenser fins 36 to reduce thermalcontact resistance and thus increase thermal conductivity between theconvective fins 18 and the condenser fins 36. In such an embodiment, thehermetically sealed chamber 16 and the convective fins 18 may be madefrom a single piece of material such as, for example, copper, aluminiumor other highly thermally conductive material that is compatible withthe working fluid 24, via a moulding or machining process.

As shown in FIG. 1B, the sleeve 20 is provided with threaded holes 44 toreceive screws (not shown) for securing the attachment member 22 to theevaporator section 34.

The attachment member 22 provides an interface between the hermeticallysealed chamber 16 and the electronic component. As can be seen fromFIGS. 1B and 2B, the attachment member 22 is provided with firstthreaded holes 46 corresponding to the threaded holes 44 on the sleeve20 and second threaded holes 48 for attachment, for example, to amotherboard of a computer. The attachment member 22 is secured to thesleeve 20 by passing screws (not shown) through respective pairs of thethreaded holes 44 on the sleeve 20 and the first threaded holes 46 onthe attachment member 22, and fastening the screws. The attachmentmember 22 is similarly secured, for example, to the motherboard of thecomputer. The leg support 22 may be of any configuration depending onthe layout of the surface to which the cooling device 10 is to besecured.

The operation of the cooling device 10 will now be described withreference to FIG. 3 which shows a schematic cross-sectional view of thecooling device 10 taken along section lines A-B and B-C in FIG. 2B.Accordingly, the left side of FIG. 3 shows the air flow through theconvective fins 18 of the cooling device 10 and the right sideillustrates circulation of the working fluid 24 within the hermeticallysealed chamber 16.

When in use, an electronic component generates and dissipates heat. Theheat from the electronic component is taken in at the base portion ofthe evaporator section 34 and raises the temperature of the base portionof the evaporator section 34.

Boiling of the working fluid 24 inside the evaporator section 34 of thehermetically sealed chamber 16 occurs as heat is taken from the baseportion of the evaporator section 34 by the working fluid 24 as latentheat of vaporization. The capillary structure 40 in the evaporatorsection 34 enhances boiling of the working fluid 24 as it creates amicro-flow situation within the evaporator section 34 by drawing theheated working fluid 24 away from the base portion of the evaporatorsection 34 and fresh working fluid 24 in the form of condensate 38 fromthe condenser fins 36 to the evaporator section 34 (see arrow 50 in FIG.3). The degree of superheat (i.e. the difference between the temperatureof the base portion of the evaporator section 34 and the saturationtemperature of the working fluid 24) is thus reduced.

Evaporation of the working fluid 24 in the capillary structure 40 andthe escape of vapour bubbles from the liquid bulk in the evaporatorsection 34 to the liquid-vapour interface result in the creation of ahigher pressure region in the evaporator section 34, forcing the vapourin the evaporator section 34 to a relatively lower pressure region inthe hollow condenser fins 36 (see arrows 52 in FIG. 3).

Condensation occurs on inner walls 54 of the condenser fins 36 as theinner walls 54 are at a lower temperature than the vaporised workingfluid 24 in the condenser fins 36. Accordingly, the working fluid 24 iscollected as condensate 38 on the inner walls 54 of the condenser fins36. Depending on the orientation of the cooling device 10, thecondensate 38 may be returned to the evaporator section 34 bygravitational means where the condenser fins 36 are at a higherelevation than the evaporator section 34, by capillary action via thecapillary element 42 (particularly when the condenser fins 36 are at alower elevation than the evaporator section 34) or a combination ofboth.

Heat in the form of latent heat of condensation transmitted from theworking fluid 24 to the condenser fins 36 is conducted to the convectivefins 18 which provide a large surface area for heat dissipation. Heat isconvected away from the convective fins 18 to ambient space by the flowof air from the fan 12 (see arrows 56 in FIG. 3). More particularly, thefan 12 mounted at the top of the main body of the cooling device 10 andlocated inside the opening end of the air guide 14 forces air to flowthrough the array of convective fins 18. The air guide 14 constricts andguides the air flow into the array of convective fins 18.

Although illustrated in an upright position in FIGS. 1A, 1B, 2A, 2B and3, it should be understood that the cooling device of the presentinvention is not limited to a particular orientation. Rather, thecooling device may be used in any orientation as long as the capillarystructure is in contact with the bulk working fluid in liquid phase forcapillary pumping.

Simulation Results

A numerical simulation was carried out to gauge the performance of thecooling device 10 using water as the working fluid 24 and copper for thechamber 16 and the convective fins 18. Convective cooling is provided byan axial flow fan 12 with a diameter of 92 millimetres (mm) and whichgenerates an air flow rate of about 11.80 litre per second (l/s)(equivalent to 25 cubic feet per minute (CFPM)). The governing equationsfor the simulation (i.e. conservation equations for energy and massprovided below) are solved simultaneously using the InternationalMathematical and Statistical Library (IMSL) running on a FortranPowerStation platform where the convergence criterion is a tolerance of10⁻⁹. The steady state performance parameters of the cooling device 10subjected to 10 watts per square centimetre (W/cm²) of heat flux from aCPU measuring 3 by 3 square centimetre (cm²) at ambient air temperatureof 30 degrees Celsius (° C.) are shown in Table 1 below.

TABLE 1 Thickness No. of CPU Air Flow of Conv. Conv. TemperatureR_(conv) R_(total) Rate P_(drop) Fins (mm) Fins (° C.) (K/W) (K/W) (l/s)(Pa) 0.15 18 49.1 0.123 0.218 11.89 24.99 0.25 17 48.1 0.116 0.206 11.9424.91 0.50 14 48.7 0.123 0.213 10.48 26.71 0.75 12 49.9 0.136 0.226 9.3428.18 1.00 10 51.2 0.151 0.241 9.25 28.29wherein R_(conv) represents the convective thermal resistance betweenthe convective fins 18 and ambient air measured in kelvin per watt(K/W), R_(total) represents the total thermal resistance of the coolingdevice 10 and P_(drop) represents the pressure drop across theconvective fins 18 and is measured in pascal (Pa).

The optimal performance for convective fins 18 of various thicknesses isshown in Table 1. As can be seen from Table 1, the cooling device 10with the thinnest convective fins 18 measuring 0.25 mm outperforms allothers by maintaining the CPU temperature at a mere 48.1° C. with anoverall but optimal thermal resistance of 0.206 K/W as it allowsplacement of an optimum number of convective fins 18 while allowing anadequate amount air flow from the axial flow fan 12 and thus achievesthe highest convective heat transfer. However, as convective finthickness is decreased further, it was observed that the benefits ofincreasing convective surface area by reducing fin thickness is somewhatoffset by a drop in the fin efficiency for the thinner fins. It isconcluded therefore that fin strength is also a design consideration.

Referring now to FIG. 4, the transient performance of the cooling device10 with convective fin thickness of 0.25 mm is shown. The data for thegraph of FIG. 4 was obtained by performing a simulation test on acooling device 10 with seventeen (17) convective fins 18, eachconvective fin having a thickness of 0.25 mm, a height of 30 mm and alength of 40 mm, the cooling device 10 being subjected to varying heatflux loads ranging from 10 W/cm² to 25 W/cm² from a 9 cm² CPU footprint.Temperatures of the CPU, the base portion of the evaporator section 34,the saturated working fluid 24 in the hermetically sealed chamber 16,and the inner walls 54 of the condenser fins 36 (in particular, at thebase portion of the condenser fin array) were measured and plottedagainst time till a steady state was reached.

In all the simulated cases, the CPU chip temperature did not exceed 70°C.—the thermal design limit of the CPU—even at a CPU heat flux of 25W/cm² (i.e. a total of 225 watts (W)).

The following were observed from the results of the simulation:

(i) The thermal resistance between the convective fins 18 and ambientair is highest for all simulated heat flux loads.

(ii) The next highest thermal resistance in the cooling device 10 is thesuperheat for boiling. In this regard, it was observed that the heattransfer coefficient of boiling increases with increasing surface heatflux at the base portion of the evaporator section 34. Thus, thecorresponding increase in boiling superheat is minimal.

(iii) The temperature differential between the CPU and the base portionof the evaporator section 34 was observed to increase from a smalldifferential of 2.3° C. to a high of 5.8° C. This indicates that boththe heat spreading and interface resistances increase with increasingheat flux loads.

(iv) The temperature differential between the condenser wall and thesaturated working fluid 24 was observed to increase from a smalldifferential of 0.5° C. to 1.5° C. This small increase is acceptable andindicates that condensation is an efficient process for heat rejection.

Observations (i) and (ii) are consistent with existing case studies.

The heat spreading problem may be alleviated by using an evaporator witha smaller base area. In regard to the boiling resistance, the thermalconductivity of the working fluid 24 may be enhanced by addingnanoparticles to the working fluid 24 to reduce the boiling superheat.The size of the convective fins 18 may be varied according to the spaceavailable to accommodate future higher heat fluxes.

Equations for Simulation Energy Balance at Base Portion of EvaporatorSection

The energy balance at the base portion of the evaporator section 34 madeof copper, taking into consideration the heat given out by the CPU chipand the heat extracted by evaporation of the working fluid 24, may bemodelled with the following equation:

$\begin{matrix}{{\lbrack ( {M\; c_{p}} )^{base} \rbrack \frac{T^{base}}{t}} = {{{- q_{flux}^{evap}}A^{evap}} + {q_{flux}^{chip}A^{chip}}}} & (1)\end{matrix}$

wherein (Mc_(p))^(base) represents the thermal capacity of the baseportion of the evaporator section 34,

$\frac{T^{base}}{t}$

represents the rate of change of the temperature at the base portion ofthe evaporator section 34 over time, q_(flux) ^(chip) represents theheat flux input from the heat source, A^(chip) represents the footprintarea of the heat source, A^(evap) represents the bottom surface area ofthe evaporator section 34, and q_(flux) ^(evap) represents the heattaken in by the working fluid 24 for evaporation.

Energy Balance of Working Fluid in Hermetically Sealed Chamber

The hermetically sealed chamber 16 includes a porous structure 40 and asaturated working fluid 24. Latent heat is taken in by the working fluid24 from the superheated surface of the base portion of the evaporatorsection 34, and is transported and rejected to the finned condensersection. The energy balance of the working fluid 24 may be modelled withthe following equation:

$\begin{matrix}{{\lbrack ( {M\; c_{p}} )^{eff} \rbrack \frac{T^{chamber}}{t}} = {{{- {h_{g}( T^{evap} )}}\frac{m_{vap}}{t}} + {{h_{f}( T^{cond} )}\frac{m_{liq}}{t}}}} & (2)\end{matrix}$

wherein (Mc_(p))^(eff) represents the effective thermal capacity of theworking fluid 24 and is obtained by summing the thermal capacity of theworking fluid 24 in the vapour, liquid and phases (i.e.(Mc_(p))^(eff)=(MCp)^(vapor)+(MCp)^(liquid)+(MCp)^(foam)),

$\frac{T^{chamber}}{t}$

represents the rate of change of the temperature of the vapour space inthe chamber 16 over time, h_(g) and h_(f) represent the respectiveenthalpies of the working fluid 24 in vapor and liquid phases at a giventemperature, T_(evap) represents the temperature of the section 34,T^(cond) presents the temperature of the condenser section,

$\frac{m_{vap}}{t}$

represents the mass flow rate of the working fluid 24 in vapour phaseand

$\frac{m_{liq}}{t}$

represents the mass flow rate of the condensate 38.

The Rohsenow Correlation, adjusted for sub-atmospheric conditions asshown in equations (3) and (4) below, is used to predict the heat fluxtaken in by the working fluid 24 for boiling at a certain superheat atsub-atmospheric conditions in the presence of the porous structure 40:

$\begin{matrix}{T^{surface} = {T^{evap} + {\overset{\overset{{Rohsenow}\mspace{14mu} {Correlations}}{}}{{( \frac{C_{sf}h_{fg}\Pr_{l}^{s}}{c_{pf}} )\lbrack {\frac{q_{flux}^{evap}}{\mu_{l}h_{fg}}\sqrt{\frac{\sigma}{g( {\rho_{l} - \rho_{g}} )}}} \rbrack}^{0.33}}( \frac{P}{P_{{at}\; m}} )^{m}( \frac{A_{wetted}}{A_{base}} )^{n}}}} & (3) \\{q_{flux}^{evap} = {( \frac{\mu_{l}h_{fg}}{\sqrt{\frac{\sigma}{g( {\rho_{l} - \rho_{g}} )}}} )\begin{bmatrix}{( {T^{surface} - T^{evap}} )( \frac{c_{pf}}{C_{sf}h_{fg}\Pr_{l}^{s}} )} \\{( \frac{P_{{at}\; m}}{P} )^{m}( \frac{A_{base}}{A_{wetted}} )^{n}}\end{bmatrix}}^{({1/0.33})}} & (4)\end{matrix}$

wherein s is 1, m is 0.293, n is −0.0984, T^(surface) represents thetemperature of the base portion of the evaporator section 34, C_(Sf), anempirical constant related to the fluid-heater surface combination, is0.0132, h_(fg) represents the latent heat of evaporation of the workingfluid 24, Pr_(l) represents the Prantl Number, C_(pf) represents theheat capacity of the working fluid 24 in liquid phase, μ_(l) representsthe viscosity of the working fluid 24 in liquid phase, σ representssurface tension, g represents acceleration due to gravity, ρ_(l)represents the density of the working fluid 24 in liquid phase, ρ_(g)represents the density of the working fluid 24 in vapour phase, Prepresents the pressure in the evaporator section 34, P_(atm) representsatmospheric pressure, A_(wetted) represents the wetted surface area ofthe porous structure 40 and A_(base) represents the surface area of thebase portion of the evaporator section 34.

The Thin Film Condensation Correlation is used to predict the heattransfer coefficient h_(i) of the condenser section as shown in equation(5) below on the assumption that the walls of the condenser section areacting as individual plates:

$\begin{matrix}{h_{i} = {0.934\lbrack \frac{{\rho_{l}( {\rho_{l} - \rho_{g}} )}{gh}_{fg}k^{3}}{{\mu ( {T_{v} - T_{cond}} )}L_{fin}} \rbrack}} & (5)\end{matrix}$

wherein k represents the thermal conductivity of the working fluid 24 inliquid phase, μ represents the viscosity of the working fluid 24 inliquid phase, T_(v) represents the temperature of the working fluid 24in vapour phase in the hermetically sealed chamber 16 and L_(fin)represents the length of the condensate 38 flow path.

Mass Balance

Mass balance was performed on the hermetically sealed chamber 16 toaccount for all the liquid and vapour mass inside the chamber 16. Thisprovides information for determining the optimum amount of working fluid24 to fill the hermetically sealed chamber 16 with to prevent drying outand to ensure wetting of the capillary structure 40 throughout theoperation of the cooling device 10. The mass balance was modelled withthe following equations:

$\begin{matrix}{\frac{M^{vapor}}{t} = {{\overset{.}{m}}_{evaporation} - {\overset{.}{m}}_{condensation}}} & (6) \\{\frac{M^{liquid}}{t} = {{\overset{.}{m}}_{condensation} - {\overset{.}{m}}_{evaporation}}} & (7)\end{matrix}$

wherein

$\frac{M^{vapor}}{t}$

represents the net flow rate of the vapour in the chamber 16,

$\frac{M^{liquid}}{t}$

represents the net flow rate of the liquid in the chamber 16, in {dotover (m)}_(evaporation) represents the rate of evaporation and {dot over(m)}_(condensation) represents the rate of condensation.

The rate of evaporation {dot over (m)}_(condensation) and the rate ofcondensation {dot over (m)}_(condensation) may be determined with thefollowing equations:

$\begin{matrix}{{\overset{.}{m}}_{evaporation} = \frac{Q_{flux}^{evap}}{h_{fg}( T_{chamber} )}} & (8) \\{{\overset{.}{m}}_{condensation} = \frac{{h_{i}( {T^{chamber} - T^{cond}} )}A_{cond}}{h_{fg}( T^{chamber} )}} & (9)\end{matrix}$

wherein Q_(flux) ^(evap) represents the heat flux at the base portion ofthe evaporator section 34 and is equivalent to q_(flux) ^(evap) definedabove, T_(chamber) represents the temperature in the vapour space of thehermetically sealed chamber 16 and A_(cond) represents the surface areaof the inner walls 54 of the condenser fins 36.

Energy Balance at Condenser Section

An energy balance was performed on the condenser fins 36. Latent heat istransferred to the inner walls 54 of the condenser fins 36 when theworking fluid 24 condenses on the inner walls 54. The heat transmittedto the condenser fins 36 is then conducted to the convective fins 18where it is transferred to ambient air by forced convection. The energybalance at the condenser section may be modelled with the followingequation:

$\begin{matrix}{{\lbrack ( {M\; c_{p}} )^{cond} \rbrack \frac{T^{cond}}{t}} = {{{\overset{.}{m}}_{condensation}{h_{fg}( T^{chamber} )}} - \frac{( {T_{cond} - T_{air}} )}{R_{hs}}}} & (10)\end{matrix}$

wherein (Mc_(p))^(cond) represents the heat capacity of the condensersection,

$\frac{T^{cond}}{t}$

represents the rate of change of the temperature of the inner walls 54of the condenser fins 36 over time, T_(air) represents ambienttemperature, and R_(hs) represents the thermal resistance between theconvective fins 18 and ambient air.

As is evident from the foregoing discussion, the present inventionprovides a compact cooling device with reduced heat spreading resistancefor intense heat emitting surfaces such as, for example, amicroprocessor or CPU of a desktop computer, as well as surfaces withhigh localised heat fluxes. The cooling device includes: an evaporatorsection where working fluid evaporates at its saturation pressure,extracting its latent heat of vaporization from the intense heatsurface; a condenser section to reject latent heat from incoming workingfluid in vapour phase from the evaporator section; a capillary structurelocated in the evaporator section to circulate the working fluid inliquid phase from the condenser section back to the evaporator section;an array of convective fins as a means of extending the heat sinksurface area; and a fan to blow air through the array of convectivefins, thereby enhancing heat rejection to ambient air. Advantageously,the evaporator section, the condenser section, the convective fins, andthe fan can all be integrated into one compact device. By using atwo-phase system, the overall size of the cooling device and heatspreading resistance can be significantly reduced, increasing theoverall heat transfer capacity of cooling device.

While the preferred embodiments of the invention have been illustratedand described, it will be clear that the invention is not limited tothese embodiments only. Numerous modifications, changes, variations,substitutions and equivalents will be apparent to those skilled in theart without departing from the scope of the invention as described inthe claims.

Further, unless the context dearly requires otherwise, throughout thedescription and the claims, the words “comprise”, “comprising” and thelike are to be construed in an inclusive as opposed to an exclusive orexhaustive sense; that is to say, in the sense of “including, but notlimited to”.

1. A cooling device, comprising: a chamber having an evaporator sectionand a condenser section, the condenser section extending from aperiphery of the evaporator section.
 2. The cooling device according toclaim 1, wherein the condenser section comprises a plurality ofcondenser fins extending from the periphery of the evaporator section.3. The cooling device according to claim 2, further comprising aplurality of convective fins extending from surfaces of the condenserfins.
 4. The cooling device according to claim 3, wherein the convectivefins extend between adjacent ones of the condenser fins.
 5. The coolingdevice according to claim 4, further comprising a fan arranged to directa flow of air at the convective fins.
 6. The cooling device according toclaim 5, further comprising an air guide arranged to direct the flow ofair from the fan to the convective fins.
 7. The cooling device accordingto claim 3, wherein the convective fins are integrally moulded to thesurfaces of the condenser fins.
 8. The cooling device according to claim1, further comprising a capillary element arranged to facilitate returnof condensate from the condenser section to the evaporator section. 9.The cooling device according to claim 1, wherein a base portion of thecondenser section is at an angle to a base portion of the evaporatorsection.
 10. The cooling device according to claim 1, wherein theevaporator section has a capillary structure.
 11. The cooling deviceaccording to claim 10, wherein the capillary structure comprises an opencell foam metal.
 12. The cooling device according to claim 11, whereinthe open cell foam metal is bonded to a base portion of the evaporatorsection by diffused bonding or brazing.
 13. The cooling device accordingto claim 10, wherein the capillary structure has a pore density of fromabout 8 to about 52 pores per centimetre (ppc).
 14. The cooling deviceaccording to claim 13, wherein the pore density of the capillarystructure is about 16 ppc.
 15. The cooling device according to claim 1,further comprising a working fluid arranged to absorb and transfer heatenergy.
 16. The cooling device according to claim 15, wherein theworking fluid has a latent heat of vaporisation of greater than or equalto about 1500 joules per gram (J/g).
 17. The cooling device according toclaim 15, further comprising a plurality of nanoparticles in workingfluid, the nanoparticles having a thermal conductivity of between about200 Watt per metre Kelvin (W/m·K) to about 400 W/m·K.